Bearing systems for small turbochargers have gone through years of development to overcome initial problems that impeded their achieving enough durability to become a commercial success. Turbochargers small enough to be applied to diesel truck and bus engines first appeared in the late 1940's in the U.S. The Elliott Company in Jeanette, Pa., designed and developed one of the first small models that went into limited production using conventional sleeve journal bearings that had an attached radial flange to carry rotor thrust.
Diesel engines at that time were just beginning to be turbocharged, and early turbochargers were designed to be stiff shaft machines where their maximum rotational speed was below the first critical speed of the rotating assembly. As diesel engines were improved structurally to allow higher power output ratings, it became necessary to develop small turbochargers that were capable of providing higher charge air pressure.
As the maximum rotational speed of the turbochargers was increased, the power losses in the stationary sleeve bearings became excessive and bearing and shaft diameters needed to be reduced. This led to the design of flexible shaft machines where the maximum rotational speed of the turbocharger rotor exceeded the first critical speed of the rotor assembly. Since the rotor assemblies then had to run through their first critical speed, bearing systems needed to be devised that would damp the amplitude of resonant vibration of the rotor assembly as it passed through its first critical speed.
In the course of bearing system development, a phenomenon termed “shaft whirl”, or “oil-film whirl”, appeared in certain bearing system designs that caused unsatisfactory performance and early bearing failures occurred. This “shaft-whirl” phenomenon is described in “Mechanical Vibrations” by Den Hartog, 2nd edition, published in 1940 by McGraw Hill. To fully investigate this phenomenon, a set of electronic equipment was invented in the late 1950's, consisting of magnetic proximity pickups mounted close to an extension of the rotor shaft that were capable of measuring the amplitude of shaft orbiting as the turbocharger was operated through its entire speed range. The results of the use of this equipment is illustrated in U.S. Pat. No. 3,056,634, where FIG. 8 shows turbocharger shaft orbiting patterns up to a speed of 78,000 RPM where the phenomenon of “shaft whirl” occurred. In FIG. 9 of the cited patent, shaft-orbiting patterns are illustrated wherein very stable operation of the turbocharger rotor was achieved by the use of a unique bearing system that suppressed the excessive orbital motion of the sleeve bearing design.
Subsequently designed bearing systems that damped the amplitude of resonant vibrations and suppressed the phenomenon of “oil whirl” are shown in U.S. Pat. No. 3,096,126 and U.S. Pat. No. 3,390,926. Both of these patents show successful bearing systems that have been used extensively in commercial turbochargers. The underlying principle leading to this success of these systems is the use of floating sleeve bearings that have an inner and outer oil film, and the bearing is allowed to rotate at a fraction of the speed of the shaft. The rotation of the bearings reduces the bearing friction loss and the inner and outer oil films allow the rotor assembly of the turbocharger to find and rotate about its center of mass. The two oil films provide sufficient damping to limit the amplitude of resonant vibrations of the rotor as it passes through its first critical speed.
The sleeve bearing systems described thus far need a separate stationary thrust bearing to carry the axial thrust loads of the turbocharger rotating assembly. A collar is provided on the shaft to bear against a stationary thrust bearing, and the high rotational speed of the collar relative to the stationary thrust bearing results in a high friction loss which, in addition to the friction losses in the sleeve bearings, results in a substantial total friction loss for the complete bearing system. This bearing system friction loss is a detriment to rapid acceleration of the turbocharger rotor.
Since it is highly advisable to have a bearing system that has a very high mechanical efficiency, an extensive effort has been expended to design and develop systems that can take advantage of the low friction losses in anti-friction or ball bearings. This effort has been recently successful, as illustrated in U.S. Pat. No. 7,677,041 B2, where angular contact ball bearings are mounted in a rotatable carrier and allow very rapid acceleration of the turbocharger rotor due to extremely low friction losses. This system is now being successfully used in commercial turbochargers.
The ball bearing system referred to above requires a supply of pressurized lubricating oil taken from the engine lube oil system. The use of lube oil in turbochargers has given rise to a number of problems over the years. Piston ring seals are used in commercial turbochargers to prevent oil leakage into the compressor casing and turbine casing. These seals are not of the positive type and a slight amount of oil can leak past the small clearance around the piston rings during certain engine operating conditions, i.e. at low idle speed. Oil leakage into the compressor casing gets carried into the engine air intake system and gets burned with the fuel in the engine cylinders, contributing to increased exhaust emissions. Oil leakage into the turbine casing mixes with the exhaust gas, also causing an increase in exhaust emissions.
In cold weather, there can be a significant time lag before viscous lube oil reaches the turbocharger bearings when the engine is started. In extreme cold weather, this time lag can cause bearing failure due to lack of oil.
Also, a hot shutdown of an engine after being operated at full load can cause residual oil in the bearing system to carbonize. Repeated hot shutdowns of an engine can eventually cause turbocharger bearings to fail due to hard carbon build-up in the bearing housing. In addition, there are limitations to where and how the turbocharger is mounted on the engine due to the necessity of providing gravity drainage of the lube oil from the bearing housing back to the engine crankcase. The cost of lube oil piping to and from the turbocharger is another consideration.
All the above-named problems associated with lube oil in turbochargers are eliminated by devising a bearing system that does not use an oil supply from the engine on which it is mounted. U.S. Pat. No. 7,025,579 B2 discloses a bearing system that uses no lube oil and is currently in commercial production. The invention described herein is also an oil-less bearing system and represents a significant improvement over the bearing system shown in U.S. Pat. No. 7,025,579 B2.
The turbocharger bearing system illustrated in U.S. Pat. No. 7,025,579 B2 uses grease-lubricated ball bearings mounted in a cylindrical bearing carrier which is carried by “O” rings between the bearing housing bore and the outside diameter of the bearing carrier. The “O” rings prevent rotation of the bearing carrier but still allow the rotor assembly of the turbocharger to find and rotate about its center of mass. The cylindrical bearing carrier is provided with a radially extending flange on one end to carry rotor thrust transmitted through the angular contact bearings to the stationary housing members. This flange must be supplied with a lubricating fluid to prevent wear and fretting of the mating surfaces of the flange and stationary housing members on which it bears to carry thrust.
As shown in U.S. Pat. No. 7,025,579 B2, a cooling jacket is provided in the bearing housing fed by liquid coolant from the engine cooling system. The outside diameter of the cylindrical bearing carrier is exposed to the coolant over a portion of its length to carry away heat generated in the bearings. This bearing system eliminates all the problems and cost associated with lubricating oil being introduced into the turbocharger bearing housing and, subsequently, carried back to the engine crank case. Another advantage of a bearing system that does not use oil is complete flexibility as to where and how the turbocharger can be mounted on the engine.
The invention described herein represents a significant improvement over the bearing system described in U.S. Pat. No. 7,025,579 B2 by providing a means of eliminating the radially extending flange on the end of the bearing carrier and eliminating the corresponding mating thrust faces on the stationary housing parts, thus significantly reducing the cost of the system.